EP0881335B1 - Engine control system for construction machine - Google Patents

Engine control system for construction machine Download PDF

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Publication number
EP0881335B1
EP0881335B1 EP98109533A EP98109533A EP0881335B1 EP 0881335 B1 EP0881335 B1 EP 0881335B1 EP 98109533 A EP98109533 A EP 98109533A EP 98109533 A EP98109533 A EP 98109533A EP 0881335 B1 EP0881335 B1 EP 0881335B1
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EP
European Patent Office
Prior art keywords
engine
injection timing
hydraulic pump
fuel injection
load
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Expired - Lifetime
Application number
EP98109533A
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German (de)
English (en)
French (fr)
Other versions
EP0881335A3 (en
EP0881335A2 (en
Inventor
Kazunori Nakamura
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Hitachi Construction Machinery Co Ltd
Original Assignee
Hitachi Construction Machinery Co Ltd
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Filing date
Publication date
Application filed by Hitachi Construction Machinery Co Ltd filed Critical Hitachi Construction Machinery Co Ltd
Publication of EP0881335A2 publication Critical patent/EP0881335A2/en
Publication of EP0881335A3 publication Critical patent/EP0881335A3/en
Application granted granted Critical
Publication of EP0881335B1 publication Critical patent/EP0881335B1/en
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Classifications

    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2221—Control of flow rate; Load sensing arrangements
    • E02F9/2232—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps
    • E02F9/2235—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps including an electronic controller
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2246—Control of prime movers, e.g. depending on the hydraulic load of work tools
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2278—Hydraulic circuits
    • E02F9/2285—Pilot-operated systems
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2278—Hydraulic circuits
    • E02F9/2292—Systems with two or more pumps
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2278—Hydraulic circuits
    • E02F9/2296—Systems with a variable displacement pump

Definitions

  • the present invention relates to an engine control system for a construction machine, and more particularly to an engine control system for a construction machine such as a hydraulic excavator wherein a diesel engine having an electronic fuel injection device (electronic control governor) is used as a prime mover.
  • a diesel engine having an electronic fuel injection device electroactive control governor
  • a construction machine such as a hydraulic excavator generally includes at least one hydraulic pump for driving a plurality of actuators, and a diesel engine is used as a prime mover for rotatively driving the hydraulic pump.
  • the diesel engine is controlled in injected fuel amount and fuel injection timing by a fuel injection device.
  • the fuel injection timing has been conventionally determined by a mechanical timer mechanism depending on a revolution speed in most cases.
  • the fuel injection timing has become freely controllable by an injection timing control actuator in addition to the injected fuel amount.
  • good combustion is realized and engine performance is improved in a wide range by determining the optimum injection timing depending on a status variable such as the engine revolution.
  • JP, A, 1-110839 discloses an internal combustion engine with a turbocharger wherein an intake pressure is detected by a pressure sensor at the time of quick acceleration to control the fuel injection timing such that the timing is advanced a predetermined angle to reduce the generation of black smoke when the detected intake pressure is not higher than a setting reference value, and is not advanced to prevent an abnormal rise of pressure in a cylinder when the detected intake pressure is not lower than the setting reference value.
  • Figs. 1 and 2 of the Publication show that an engine load is input as one item of information to be reflected in control of the injection timing.
  • Other solutions are disclosed in US-A-5 447 138 and US-A-5 218 945 but without load dependancy.
  • the conventional electronic fuel injection device for a diesel engine has intended to realize combustion with a less amount of No x , etc. by adjusting the fuel injection timing depending on an engine load.
  • the engine load is estimated from an engine revolution speed and an injected fuel amount, and is not accurately detected in a direct manner. This has raised the problem that the fuel injection timing cannot be controlled with high accuracy and there is a limit in effect of improving combustion.
  • an object to be driven by the engine is a hydraulic pump.
  • a delivery rate and a delivery pressure of the hydraulic pump are frequently changed and a load of the hydraulic pump, i.e., an engine load, is fluctuated.
  • injection timing control is performed by estimating the load based on the engine revolution speed and the injected fuel amount in such a diesel engine, in particular, the injection timing cannot be controlled with good response following fluctuation in load of the hydraulic pump and a sufficient improvement of combustion cannot be obtained.
  • An object of the present invention is to provide an engine control system for a construction machine with which, in a diesel engine for rotatively driving a hydraulic pump, combustion is improved and engine performance is enhanced by controlling the fuel injection timing with good response and high accuracy following load fluctuation.
  • the above-stated fuel injection timing control based on the engine load can be performed in combination with fuel injection timing control based on the revolution speed.
  • Fig. 1 is a diagram showing an entire configuration of an engine control system according to a first embodiment of the present invention along with a hydraulic circuit and a pump control system.
  • Fig. 2 is an enlarged view of a regulator section of a hydraulic pump.
  • Fig. 3 is a diagram showing a schematic configuration of an electronic fuel injection device.
  • Fig. 4 is a functional block diagram showing a sequence of processing steps in a pump controller.
  • Fig. 5 is a functional block diagram showing a sequence of processing steps in an engine controller.
  • Fig. 6 is a functional block diagram showing a sequence of processing steps in a fuel injection timing calculation block in the engine controller.
  • Fig. 7 is a graph showing the relationship among an engine revolution speed, an engine load and injection timing resulted under control made by the engine control system of the present invention.
  • Fig. 8 is a diagram showing an entire configuration of an engine control system according to a second embodiment of the present invention along with a hydraulic circuit-and a pump control system.
  • Fig. 9 is a functional block diagram showing a sequence of processing steps in a pump controller.
  • reference numerals 1 and 2 denote variable displacement hydraulic pumps.
  • the hydraulic pumps 1, 2 are connected to actuators 5, 6 through valve units 3, 4, respectively, and the actuators 5, 6 are driven by hydraulic fluids delivered from the hydraulic pumps 1, 2.
  • the actuators 5, 6 are hydraulic cylinders for, e.g., moving a boom, an arm, etc. which constitute a working front of a hydraulic excavator, and predetermined work is performed with driving of the actuators 5, 6.
  • Commands for driving the actuators 5, 6 are applied from control lever units 33, 34 and the valve units 3, 4 are operated upon the control lever units 33, 34 being manipulated.
  • the hydraulic pumps 1, 2 are, by way of example, swash plate pumps wherein tiltings of swash plates 1a, 1b serving as displacement varying mechanisms are controlled by regulators 7, 8 to control respective pump delivery rates.
  • Denoted by 9 is a fixed displacement pilot pump serving as a pilot pressure generating source which generates a hydraulic pressure signal and a hydraulic fluid for control.
  • the hydraulic pumps 1, 2 and the pilot pump 9 are coupled to an output shaft 11 of a prime mover 10 and are rotatively driven by the prime mover 10.
  • the prime mover 10 is a diesel engine and includes an electronic fuel injection device 12.
  • a target revolution speed of the prime mover 10 is commanded by an accelerator operation input unit 35.
  • the regulators 7, 8 of the hydraulic pumps 1, 2 comprise, respectively, tilting actuators 20, 20, first servo valves 21, 21 for positive tilting control, and second servo valves 22, 22 for input torque limiting control.
  • the servo valves 21, 22 control hydraulic fluid pressures acting on the tilting actuators 20 from the pilot pump 9.
  • the regulators 7, 8 of the hydraulic pumps 1, 2 are shown in Fig. 2 in an enlarged scale.
  • the tilting actuators 20 each comprise an operating piston 20c provided with a large-diameter pressure bearing portion 20a and a small-diameter pressure bearing portion 20b at opposite ends thereof, and pressure bearing chambers 20d, 20e in which the pressure bearing portions 20a, 20b are positioned respectively.
  • the operating piston 20c is moved to the right on the drawing due to an area difference between the pressure bearing portions 20a, 20b, whereupon the tilting of the swash plate 1a or 2a is diminished to reduce the pump delivery rate.
  • the operating piston 20c When the pressure in the pressure bearing chamber 20d on the large-diameter side lowers, the operating piston 20c is moved to the left on the drawing, whereupon the tilting of the swash plate 1a or 2a is enlarged to increase the pump delivery rate. Further, the pressure bearing chamber 20d on the large-diameter side is connected to a delivery line of the pilot pump 9 through the first and second servo valves 21, 22, whereas the pressure bearing chamber 20e on the small-diameter side is directly connected to the delivery line of the pilot pump 9.
  • the first servo valves 21 for positive tilting control are each a valve operated by a control pressure from a solenoid control valve 30 or 31.
  • a valve body 21a When the control pressure is high, a valve body 21a is moved to the right on the drawing, causing a pilot pressure from the pilot pump 9 to be transmitted to the pressure bearing chamber 20d without being reduced, whereby the delivery rate of the hydraulic pump 1 or 2 is reduced.
  • the valve body 21a is moved to the left on the drawing by force of a spring 21b, causing the pilot pressure from the pilot pump 9 to be transmitted to the pressure bearing chamber 20d after being reduced, whereby the delivery rate of the hydraulic pump 1 or 2 is increased.
  • the second servo valves 22 for input torque limiting control are each a valve operated by delivery pressures of the hydraulic pumps 1 and 2 and a control pressure from a solenoid control valve 32.
  • the delivery pressures of the hydraulic pumps 1 and 2 and the control pressure from the solenoid control valve 32 are introduced respectively to pressure bearing chambers 22a, 22b, 22c of operation drivers.
  • the valve body 22e As the sum of hydraulic pressure forces given by the delivery pressures of the hydraulic pumps 1 and 2 rises over the setting value, the valve body 22e is moved to the left on the drawing, causing the pilot pressure from the pilot pump 9 to be transmitted to the pressure bearing chamber 20d without being reduced, whereby the delivery rate of the hydraulic pump 1 or 2 is reduced.
  • the setting value is increased so that the delivery rate of the hydraulic pump 1 or 2 starts reducing from a relatively high delivery pressure of the hydraulic pump 1 or 2
  • the setting value is decreased so that the delivery rate of the hydraulic pump 1 or 2 starts reducing from a relatively low delivery pressure of the hydraulic pump 1 or 2.
  • the solenoid control valves 30, 31 are operated (as described later) to maximize the control pressures output from them when the control lever units 33, 34 are in neutral positions, and when the control lever units 33, 34 are manipulated, to lower the control pressures output from them with an increase in respective input amounts by which the control lever units 33, 34 are manipulated.
  • the solenoid control valve 32 is operated (as described later) to lower the control pressure output from it as the target revolution speed indicated by an accelerator signal output from the accelerator operation input unit 35.
  • the tiltings of the hydraulic pumps 1, 2 are controlled so that the delivery rates of the hydraulic pumps 1, 2 are increased to provide the delivery rates adapted for demanded flow rates of the valve units 3, 4.
  • the tiltings of the hydraulic pumps 1, 2 are controlled so that maximum values of the delivery rates of the hydraulic pumps 1, 2 are limited to smaller values to keep the load of the hydraulic pump 1 from exceeding the output torque of the prime mover 10.
  • reference numeral 40 denotes a pump controller and 50 an engine controller.
  • the pump controller 40 receives detection signals from pressure sensors 41, 42, 43, 44 and position sensors 45, 46, as well as the accelerator signal from the accelerator operation input unit 35. After executing predetermined processing, the pump controller 40 outputs control currents to the solenoid control valves 30, 31, 32 and an engine load torque signal to the engine controller 50.
  • the control lever units 33, 34 are of the hydraulic pilot type producing and outputting a pilot pressure as an operation signal.
  • Shuttle valves 36, 37 for detecting the pilot pressures are provided in respective pilot circuits of the control lever units 33, 34, and the pressure sensors 41, 42 electrically detect the respective pilot pressures detected by shuttle valves 36, 37.
  • the pressure sensors 43, 44 electrically detect the respective delivery pressures of the hydraulic pumps 1, 2, and the position sensors 45, 46 electrically detect the respective tiltings of the swash plates 1a, 2a of the hydraulic pumps 1, 2.
  • the engine controller 50 receives not only the accelerator signal from the accelerator operation input unit 35 and the engine load torque signal from the pump controller 40, but also detection signals from a revolution speed sensor 51, a link position sensor 52 and a lead angle sensor 53. After executing predetermined processing, the engine controller 50 outputs control currents to an governor actuator 54 and a timer actuator 55.
  • the revolution speed sensor 51 detects the revolution speed of the engine 10.
  • Fig. 3 shows an outline of the electronic fuel injection device 12 and a control system for it.
  • the electronic fuel injection device 12 comprises an injection pump 56, an injection nozzle 57 and a governor mechanism 58 for each cylinder of the engine 10.
  • the injection pump 56 comprises a plunger 61 and a plunger barrel 62 within which the plunger 61 is vertically movable.
  • a cam shaft 59 is rotated, a cam 60 mounted on the cam shaft 59 pushes up the plunger 61 and then pressurize fuel upon the rotation.
  • the pressurized fuel is delivered to a nozzle 57 and injected into the engine cylinder.
  • the cam shaft 59 is rotated in association with a crankshaft of the engine 10.
  • the governor mechanism 58 comprises the governor actuator 54 and a link mechanism 64 of which position is controlled by the governor actuator 54.
  • the link mechanism 64 rotates the plunger 61 to change the relationship between a thread lead of the plunger 61 and a fuel intake port formed in the plunger barrel 62, whereby an effective compression stroke of the plunger 61 is changed to adjust the injected fuel amount.
  • the link position sensor 52 is provided in the link mechanism to detect the link position.
  • the governor actuator 54 is, e.g., an electromagnetic solenoid.
  • the electronic fuel injection device 12 includes the timer actuator 55 which advances a lead angle of the cam shaft 59 with respect to rotation of a shaft 65 coupled to the crankshaft for phase adjustment to adjust the fuel injection timing. Because of necessity of transmitting a drive torque to the injection pump 56, the timer actuator 55 is required to produce large force enough for the phase adjustment. For that reason, the timer actuator 55 includes a hydraulic actuator built in it and is provided with a solenoid control valve 66 for converting the control current from the engine controller 50 into a hydraulic pressure signal and advancing the lead angle of the cam shaft 59 in a hydraulic manner.
  • the revolution speed sensor 51 is provided to detect a revolution speed of the shaft 65 and the lead angle sensor 53 is provided to detect a revolution speed of the cam shaft 69.
  • Fig. 4 shows a sequence of processing steps in the pump controller 40 in the form of a functional block diagram.
  • the detection signals (pilot lever sensor signals P1 and P2) from the pressure sensors 41, 42 are converted into target tiltings ⁇ 01 , ⁇ 02 of the hydraulic pumps 1, 2 in a target tilting calculation blocks 40a, 40b and then converted into current values I 1 , I 2 in current value calculation blocks 40c, 40d.
  • Control currents corresponding to the current values I 1 , I 2 are output to the solenoid control valves 30, 31.
  • the relationships between the pilot pressures represented by the sensor signals P1, P2 and the target tiltings ⁇ 01 , ⁇ 02 in the blocks 40a, 40b are set such that as the pilot pressures rise, the target tiltings ⁇ 01 , ⁇ 02 increase.
  • the relationships between the target tiltings ⁇ 01 , ⁇ 02 and the current values I 1 , I 2 in the blocks 40c, 40d are set such that as the target tiltings ⁇ 01 , ⁇ 02 increase, the current values I 1 , I 2 increase.
  • the solenoid control valves 30, 31 are operated to maximize the control pressures output from them when the control lever units 33, 34 are in neutral positions, and when the control lever units 33, 34 are manipulated, to lower the control pressures output from them with an increase in respective input amounts by which the control lever units 33, 34 are manipulated.
  • the accelerator signal from the accelerator operation input unit 35 is converted into an allowable maximum torque T F in a maximum torque calculation block 40e and then converted into a current value I 3 in a current value converter 40f.
  • a control current corresponding to the current value I 3 is output to the solenoid control valve 32.
  • the accelerator operation input unit 35 is manipulated by an operator, and the accelerator signal is selected depending on conditions where the operator is going to use the machine, thereby commanding the target revolution speed.
  • the relationship between the accelerator signal and the allowable maximum torque T p in the block 40e is set such that the allowable maximum torque T p increases as the target revolution speed represented by the accelerator signal becomes higher.
  • the relationship between the allowable maximum torque T p and the current value I 3 in the block 40f is set such that the allowable maximum torque T p becomes larger as the current value I 3 increases.
  • the detection signal from the position sensor 45 (tilting signal ⁇ 1 of the hydraulic pump 1) and the detection signal from the pressure sensor 43 (delivery pressure signal P D1 of the hydraulic pump 1) are input to a torque calculation block 40g, while the detection signal from the position sensor 46 (tilting signal ⁇ 2 of the hydraulic pump 2) and the detection signal from the pressure sensor 44 (delivery pressure signal P D2 of the hydraulic pump 2) are input to a torque calculation block 40h.
  • the load torques T r1 , T r2 are added in an adder 40i to determine a total of the load torques of the hydraulic pumps 1, 2.
  • the total of the load torques is output as an engine load torque signal T to the engine controller 50.
  • Fig. 5 shows a sequence of processing steps in the engine controller 50 in the form of a functional block diagram.
  • the accelerator signal from the accelerator operation input unit 35, the detection signal from the revolution speed sensor 51 (engine revolution speed signal), and the detection signal from the link position sensor 52 (link position signal) are converted into an injected fuel amount command in an injected fuel amount calculation block 50a.
  • a control current corresponding to the injected fuel amount command is output to the governor actuator 54.
  • the processing executed in the injected fuel amount calculation block 50a is known.
  • the link position of the link mechanism 64 is adjusted to increase the injected fuel amount.
  • the link position of the link mechanism 64 is adjusted to reduce the injected fuel amount.
  • the link position signal is used for feedback control.
  • the detection signal from the revolution speed sensor 51 (engine revolution speed signal), the engine load torque signal T from the pump controller 40, and the detection signal from the lead angle sensor 53 (lead angle signal) are converted into a fuel injection timing command in a fuel injection timing calculation block 50b.
  • a control current corresponding to the fuel injection timing command is output to the solenoid control valve 66 of the timer actuator 55.
  • Fig. 6 shows a sequence of processing steps in the fuel injection timing calculation block 50b in more detail.
  • the detection signal from the revolution speed sensor 51 (engine revolution speed signal) is input to a first injection timing calculation block 50c where the injection timing depending on the engine revolution speed is calculated.
  • the injection timing is calculated based on the well-known concept. More specifically, the first injection timing calculation block 50c has set therein beforehand the relationship between the engine revolution speed and the injection timing with which when the engine revolution speed is low, the injection timing is relatively delayed with respect to the engine revolution and as the engine revolution speed rises, the injection timing is advanced, namely set to an earlier point in time. The injection timing is calculated from that relationship.
  • the engine load torque signal T from the pump controller 40 is input to a second injection timing calculation block 50d where the injection timing depending on the engine load torque is calculated.
  • the injection timing is calculated in the second injection timing calculation block 50d. More specifically, the second injection timing calculation block 50d has set therein beforehand the relationship between the engine load torque and the injection timing with which when the engine load torque is small, the injection timing is relatively advanced with respect to the engine revolution, and as the engine load torque increases, the injection timing is delayed. The injection timing is calculated from that relationship.
  • Values representing the injection timings calculated in the first and second injection timing calculation blocks 50c, 50d are added in an adder 50e, and a resulting total value is output as target injection timing.
  • a deviation of the target injection timing with respect to the detection signal from the lead angle sensor 53 (lead angle signal) is determined in a subtractor 50f, and based on the determined deviation, the injection timing command is calculated in a command value calculation block 50g.
  • the injection timing command is converted into a control current which is output to the solenoid control valve 66 of the timer actuator 55.
  • Fig. 7 shows the relationship among the engine revolution speed, the engine load torque and the injection timing resulted when the timer actuator 55 is controlled in accordance with the injection timing command explained above.
  • the fuel injection timing is controlled to be advanced as the engine revolution speed rises, and to be delayed as the engine load torque increases.
  • the pump controller 40 directly and accurately calculates the load imposed on the engine by calculating the load torques T r1 , T r2 of the hydraulic pumps 1, 2 and then summing up the calculated load torques to determine the engine load torque, and the engine controller 50 calculates the target fuel injection timing by using the engine load torque.
  • the target fuel injection timing depending on the engine load can be therefore determined accurately.
  • the fuel injection timing can be controlled with good response following the load fluctuation.
  • FIG. 8 and 9 A second embodiment of the present invention will be explained with reference to Figs. 8 and 9.
  • the load torque of the hydraulic pump is calculated by using a target pump tilting.
  • Figs. 8 and 9 equivalent members and functions shown in Figs. 1 and 4 are denoted by the same reference numerals.
  • a pump controller 40A receives only the detection signals from the pressure sensors 41, 42, 43, 44 and the accelerator signal from the accelerator operation input unit 35.
  • Fig. 9 shows a sequence of processing steps in the pump controller 40A in the form of a functional block diagram.
  • respective processing steps in the target tilting calculation blocks 40a, 40b, the current value calculation blocks 40c, 40d, the maximum torque calculation block 40e and the current value converter 40f are the same as in the first embodiment shown in Fig. 4.
  • the target tilting ⁇ 01 of the hydraulic pump 1 calculated in the target tilting calculation block 40a and the detection signal from the pressure sensor 43 are input to a torque calculation block 40Ag
  • the target tilting ⁇ 02 of the hydraulic pump 2 calculated in the target tilting calculation block 40b and the detection signal from the pressure sensor 44 are input to a torque calculation block 40Ah.
  • the load torques T r1 , T r2 are added in the adder 40i to determine a total T r12 of the load torques of the hydraulic pumps 1, 2.
  • the total pump load torque T r12 is input, along with the allowable maximum torque T p calculated in the maximum torque calculation block 40e, to a minimum value selection block 40j which selects smaller one of the two torques input thereto.
  • the tiltings of the hydraulic pumps 1, 2 are controlled by the regulators 7, 8 so that as the delivery pressures of the hydraulic pumps 1, 2 rise or as the target revolution speed input from the accelerator operation input unit 35 lowers, the maximum values of the delivery rates of the hydraulic pumps 1, 2 are reduced to keep the load of the hydraulic pump 1 from exceeding the output torque of the prime mover 10. More specifically, when the total load torque of the hydraulic pumps 1, 2 is going to exceed the allowable maximum torque T p in a condition where the target tiltings ⁇ 01 , ⁇ 02 of the hydraulic pumps 1, 2 calculated in the target tilting calculation blocks 40a, 40b are increased, the tiltings of the hydraulic pumps 1, 2 are controlled not to exceed the respective target tiltings at that time. Thus, by selecting smaller one of the total pump load torque T r12 and the allowable maximum torque T p in the minimum value selection block 40j, a value corresponding to the actual load torque of the hydraulic pumps 1, 2 is determined.
  • the load torque selected in the minimum value selection block 40j is output as an engine load torque signal T o to the engine controller 50.
  • the total load torque of the hydraulic pumps 1, 2 (engine load torque) is determined by using the target pump tiltings which represent values before the delivery rates of the hydraulic pumps 1, 2 are actually changed, response in injection timing control following fluctuation in the engine load caused by change in the delivery rates of the hydraulic pumps 1, 2 is further improved, the injection timing control can be performed with higher accuracy, and a further improvement of combustion can be achieved.
  • the position sensors for detecting the swash plate positions of the hydraulic pumps 1, 2 are dispensed with, the control system can be realized at a reduced cost.
  • the delivery pressures of the hydraulic pumps 1, 2 are directly detected by the pressure sensors 43, 44 in the above embodiments. However, since there is a fixed relationship between the load pressures of the hydraulic actuators 5, 6 and the delivery pressures of the hydraulic pumps 1, 2, the delivery pressures of the hydraulic pumps 1, 2 may be obtained by detecting the load pressures of the hydraulic actuators 5, 6 and estimating them from the detected load pressures.
  • the fuel injection timing of the engine is determined by calculating the accurate load imposed on the engine, the fuel injection timing can be controlled with good response and high accuracy following load fluctuation of the engine.
  • a temperature rise in the engine combustion chamber can be held down and the engine reliability can be improved.

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  • Engineering & Computer Science (AREA)
  • Mining & Mineral Resources (AREA)
  • Civil Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Structural Engineering (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Electrical Control Of Air Or Fuel Supplied To Internal-Combustion Engine (AREA)
  • Control Of Positive-Displacement Pumps (AREA)
  • Combined Controls Of Internal Combustion Engines (AREA)
  • Operation Control Of Excavators (AREA)
  • Control Of Vehicle Engines Or Engines For Specific Uses (AREA)
  • Fluid-Pressure Circuits (AREA)
EP98109533A 1997-05-27 1998-05-26 Engine control system for construction machine Expired - Lifetime EP0881335B1 (en)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
JP9137073A JPH10325347A (ja) 1997-05-27 1997-05-27 建設機械のエンジン制御装置
JP13707397 1997-05-27
JP137073/97 1997-05-27

Publications (3)

Publication Number Publication Date
EP0881335A2 EP0881335A2 (en) 1998-12-02
EP0881335A3 EP0881335A3 (en) 1999-07-21
EP0881335B1 true EP0881335B1 (en) 2003-10-15

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EP98109533A Expired - Lifetime EP0881335B1 (en) 1997-05-27 1998-05-26 Engine control system for construction machine

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US (1) US6021756A (ko)
EP (1) EP0881335B1 (ko)
JP (1) JPH10325347A (ko)
KR (1) KR100313551B1 (ko)
CN (1) CN1077648C (ko)
DE (1) DE69818907T2 (ko)

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US6341596B1 (en) * 2000-04-28 2002-01-29 General Electric Company Locomotive transient smoke control strategy using load application delay and fuel injection timing advance
US7051715B2 (en) 2002-12-03 2006-05-30 General Electriccompany Apparatus and method for suppressing diesel engine emissions
JP4484467B2 (ja) * 2003-08-01 2010-06-16 日立建機株式会社 走行式油圧作業機
US8374766B2 (en) * 2007-11-29 2013-02-12 Caterpillar Paving Products Inc. Power management system for compaction vehicles and method
US9133837B2 (en) * 2008-04-24 2015-09-15 Caterpillar Inc. Method of controlling a hydraulic system
KR101670529B1 (ko) * 2008-12-15 2016-10-31 두산인프라코어 주식회사 건설 기계의 유압펌프 유량 제어장치
US8522543B2 (en) * 2008-12-23 2013-09-03 Caterpillar Inc. Hydraulic control system utilizing feed-forward control
JP5420513B2 (ja) * 2009-12-03 2014-02-19 日立建機株式会社 油圧作業機
EP3112539B1 (en) * 2014-02-24 2017-11-22 Sumitomo (S.H.I.) Construction Machinery Co., Ltd. Shovel and shovel control method

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US6021756A (en) 2000-02-08
EP0881335A3 (en) 1999-07-21
KR19980087353A (ko) 1998-12-05
KR100313551B1 (ko) 2002-02-28
JPH10325347A (ja) 1998-12-08
CN1077648C (zh) 2002-01-09
DE69818907T2 (de) 2004-08-19
EP0881335A2 (en) 1998-12-02
DE69818907D1 (de) 2003-11-20
CN1200436A (zh) 1998-12-02

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